JP6558405B2 - Control device for compression ignition engine - Google Patents

Description

  The present disclosure relates to a control device for a compression ignition engine.

  Patent Document 1 discloses HCCI (Homogeneous-Charge Compression Ignition) execution means for executing compression self-ignition operation at least in the partial load operation region of the engine, and ignition assist when executing compression self-ignition operation in the partial load operation region. There is described a control device for a spark ignition type gasoline engine that includes an ignition assist means that executes and performs a reduced-cylinder operation in an operation region on a low load side in a partial load operation region. Furthermore, Patent Document 1 describes that combustion is performed with an excess air ratio λ = 1 during reduced-cylinder operation.

JP 2007-154859 A

  However, the conventional spark ignition type gasoline engine control device provided with the ignition assist means and the HCCI execution means adopts the form of burning with the excess air ratio λ = 1 during the reduced-cylinder operation, so that there is a problem of poor fuel consumption. There is.

  In the low-load operation region, the air-fuel ratio A / F may be leaned. However, when leaning, flame propagation during SI (Spark Ignition) combustion becomes difficult, and CI (Compression Ignition: compression) Ignition) The problem that the combustion time at the time of combustion varies also arises.

  The present disclosure solves the above-described conventional problems, and can improve fuel efficiency and stably perform SI combustion operation even during reduced-cylinder operation in a low-load operation region, and perform subsequent CI combustion operation at a desired timing. The purpose is to be able to be performed in.

  SI combustion is combustion with flame propagation that starts by forcibly igniting the air-fuel mixture in the combustion chamber. On the other hand, CI combustion is combustion that starts when the air-fuel mixture in the combustion chamber undergoes compression self-ignition. Combustion mode combining SI combustion and CI combustion means that when combustion is forcibly ignited in the air-fuel mixture in the combustion chamber and combustion by flame propagation is started, the heat generated by SI combustion and the pressure increase due to flame propagation The unburned air-fuel mixture in the combustion chamber is burned by compression ignition. Hereinafter, this combustion mode is referred to as spark ignition controlled compression ignition (SPCCI) combustion.

  In the combustion by compression ignition, when the temperature in the combustion chamber before the start of compression varies, the timing of compression ignition changes greatly. In SPCCI combustion, by adjusting the calorific value of SI combustion, it is possible to absorb temperature variations in the combustion chamber before starting compression. If the start timing of SI combustion is adjusted by adjusting the ignition timing, for example, according to the temperature in the combustion chamber before the start of compression, the timing of compression ignition can be controlled.

  In SI combustion by flame propagation, the pressure rise is more gradual than that in CI combustion, and thus SPCCI combustion can suppress the generation of combustion noise. Further, since CI combustion has a shorter combustion period than SI combustion, SPCCI combustion is advantageous in improving fuel efficiency.

  When the engine is operating in a low load region, the fuel injection amount is reduced and the temperature in the combustion chamber is also lowered. Therefore, even if trying to perform SPCCI combustion, it is difficult to ignite by ignition, and SI combustion is unstable. In addition, there is a possibility that the CI combustion cannot be performed due to insufficient temperature.

  Therefore, the inventors of the present application make it possible to realize stable SPCCI combustion while leaning the air-fuel ratio A / F even in such a low load region that is disadvantageous for SPCCI combustion.

  The technique disclosed herein provides a so-called reduced-cylinder operation in which fuel is supplied only to some of a plurality of cylinders and fuel is not supplied to the remaining cylinders under a specific condition in a low load region. The compression ratio is greater than the value.

  Specifically, the present disclosure is directed to a control device for a compression ignition engine, and the following solution is taken.

That is, the first aspect includes an engine having a plurality of cylinders, each having a combustion chamber, an ignition unit arranged in each combustion chamber, a fuel injection unit arranged facing each combustion chamber, And a control unit that outputs a control signal to each of the ignition unit and the fuel injection unit. When the engine is operated at a load higher than a predetermined load, fuel is supplied to all of the plurality of cylinders to perform all-cylinder operation, while when the engine is operated at a predetermined load or less. Supplies fuel to some of the plurality of cylinders to perform reduced-cylinder operation. Further, in the reduced-cylinder operation, the control unit controls the ignition unit so that the fuel injection unit injects fuel into a part of the plurality of cylinders and ignites the generated air-fuel mixture, With an air / fuel ratio larger than the theoretical air / fuel ratio and a high compression ratio equal to or higher than a predetermined value, the engine starts SI combustion by flame propagation by ignition of the ignition unit, and then the unburned mixture is caused by compression ignition. CI burns. Further, when the engine is operated at a predetermined load or less, the control unit is configured to perform a first combustion mode in which SI combustion and CI combustion following the SI combustion are performed by an air-fuel ratio equal to the theoretical air-fuel ratio. When the engine is operated in the first combustion mode by switching between SI combustion and CI combustion following the SI combustion with an air / fuel ratio larger than the stoichiometric air / fuel ratio, the reduced cylinder operation is not performed. .

  Here, the “engine” can be a four-stroke engine that operates when the combustion chamber repeats an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke.

According to the first aspect, in the low load region, the reduced-cylinder operation for supplying fuel to a part of the plurality of cylinders is performed, and the air-fuel ratio larger than the stoichiometric air-fuel ratio, that is, A / F lean, and SI combustion and subsequent CI combustion are performed at a high compression ratio equal to or higher than a predetermined value (as an example, the value of the geometric compression ratio is approximately 13.0 or more). Therefore, by performing the reduced-cylinder operation in the A / F lean state, discharge of unprocessed NOx (raw NOx) is suppressed and fuel efficiency is improved. In addition, since CI combustion is performed at a high compression ratio that is equal to or greater than a predetermined value during reduced-cylinder operation, variations in the combustion timing of CI combustion are less likely to occur, so that CI combustion can be performed stably, that is, at a desired timing. become able to. Further, in the first combustion mode in which the SI combustion and the CI combustion following the SI combustion are performed at an air / fuel ratio equal to the theoretical air / fuel ratio, the entire cylinder operation is performed without performing the reduced cylinder operation. The generation of NOx at the time of transition of control can be suppressed.

In a first aspect, the control unit, in the reduced-cylinder operation, may be a state of being operated also the intake valves and exhaust valves of the remainder of the plurality of cylinders.

  In this way, it is not necessary to provide an expensive valve stop mechanism. Specifically, since the air-fuel mixture is in an A / F lean state, it is not necessary to stop the valve in the cylinder that has been deactivated in order to obtain the stoichiometric air-fuel ratio, thereby reducing the pumping loss.

  In the first aspect, when the temperature of the intake air to the combustion chamber is lower than a predetermined value, the control unit may perform all-cylinder operation instead of reduced-cylinder operation.

  This eliminates the instability of SI combustion when the intake air temperature is low, so that it is possible to suppress a delay in the combustion timing of subsequent CI combustion.

  In the first aspect, when the coolant temperature of the engine is lower than a predetermined value, the control unit may perform all-cylinder operation instead of reduced-cylinder operation.

  In this way, in the all-cylinder operation, the cooling of the unburned mixture (end gas) in the vicinity of the wall surface in the combustion chamber is eliminated more quickly than in the reduced-cylinder operation, so that the combustion start timing of CI combustion is delayed. Can be suppressed.

  The first aspect may further include a swirl generator that generates a swirl flow in the combustion chamber, and the control unit may generate a swirl flow in the combustion chamber with respect to the swirl generator in the reduced-cylinder operation.

  Thus, by generating a swirl flow, SI combustion is stabilized and CI combustion is optimized. In addition, it is possible to suppress variations in torque between cycles.

  In this case, the value of the swirl ratio in the swirl flow may be 4 or more.

  In this way, as will be described later, since the swirl flow generated in the combustion chamber can be strengthened, SI combustion can be performed more stably.

  According to the control device for the compression ignition engine according to the present disclosure, it is possible to improve the fuel efficiency and perform the SI combustion operation stably even during the reduced-cylinder operation in the low load operation region, and to perform the subsequent CI combustion operation. It can be done at the timing of.

FIG. 1 is a system diagram including an engine and its accessories according to an embodiment of the present disclosure. FIG. 2 is a plan view (upper view) and a cross-sectional view (lower view) taken along line II-II showing an example of the configuration of the combustion chamber of the engine according to an embodiment of the present disclosure. FIG. 3 is a plan view illustrating an example of a configuration of a combustion chamber and an intake system of an engine according to an embodiment of the present disclosure. FIG. 4 is a block diagram illustrating an example of a configuration of an engine control device according to an embodiment of the present disclosure. FIG. 5 is a graph illustrating the engine operating region and the excess air ratio in each operating region according to an embodiment of the present disclosure (upper diagram), and a graph illustrating the engine operating region and the swirl valve opening in each operating region. (Figure below). FIG. 6 is a timing chart illustrating the fuel injection timing, the ignition timing, and the combustion waveform in each operation region of the engine according to an embodiment of the present disclosure. FIG. 7 shows the first injection timing in layer 2 (operation region (1) -2) of the engine operation region and layer 3 (operation region (1) -1) including reduced-cylinder operation according to an embodiment of the present disclosure. It is a graph which shows the relationship with load (shaft average effective pressure). FIG. 8 is a graph showing a relationship between the second injection timing and the load in the layer 2 including the engine operating region Layer 2 and the reduced cylinder operation according to the embodiment of the present disclosure. FIG. 9 is a graph showing the relationship between the third injection timing and the load in the layer 3 including the engine operating region Layer 2 and the reduced cylinder operation according to the embodiment of the present disclosure. FIG. 10 is a graph showing the relationship between the fuel amount and the first injection timing fuel amount per cycle in the layer 2 of the engine operation region and the layer 3 including the reduced cylinder operation according to an embodiment of the present disclosure. FIG. 11 is a graph showing the relationship between the fuel amount and the second injection timing fuel amount per cycle in layer 2 of the engine operation region and layer 3 including reduced-cylinder operation according to an embodiment of the present disclosure. FIG. 12 is a graph showing a relationship between the fuel amount and the third injection timing fuel amount per cycle in the layer 2 including the engine operation region according to the embodiment of the present disclosure and the layer 3 including the reduced cylinder operation. FIG. 13 is a graph showing the relationship between the total fuel amount per cycle and the load in the layer 2 including the engine operation region Layer 2 and the reduced cylinder operation according to the embodiment of the present disclosure. FIG. 14 is a graph showing the relationship between the exhaust valve closing timing and the load in the layer 3 including the layer 2 and the reduced-cylinder operation in the engine operation region according to the embodiment of the present disclosure. FIG. 15 is a graph illustrating the relationship between the intake valve opening timing and the load in the layer 3 including the engine operating region Layer 2 and the reduced cylinder operation according to the embodiment of the present disclosure. FIG. 16 is a graph showing the relationship between the amount of external EGR added and the load in layer 2 including engine 2 in the operation region and layer 3 including reduced-cylinder operation according to an embodiment of the present disclosure. FIG. 17 is a flowchart illustrating a cylinder deactivation control process of an engine according to an embodiment of the present disclosure.

(One embodiment)
An embodiment of the present invention will be described with reference to the drawings.

  Hereinafter, an embodiment of a control device for a premixed compression ignition engine will be described in detail with reference to the drawings. The following description is an example of an engine control device. FIG. 1 is a diagram illustrating the configuration of an engine. FIG. 2 is a diagram illustrating the configuration of the combustion chamber, in which the upper diagram of FIG. FIG. 3 is a diagram illustrating the configuration of the combustion chamber and the intake system. In FIG. 1, the intake side is the left side of the drawing, and the exhaust side is the right side of the drawing. 2 and 3, the intake side is the right side of the drawing, and the exhaust side is the left side of the drawing. FIG. 4 is a block diagram illustrating the configuration of the engine control device.

  The engine 1 is a four-stroke engine that operates when the combustion chamber 17 repeats an intake stroke, a compression stroke, an expansion stroke, and an exhaust stroke. The engine 1 is mounted on a four-wheeled vehicle. The vehicle travels when the engine 1 is driven. The fuel of the engine 1 is gasoline in this configuration example. The fuel may be gasoline containing bioethanol or the like. The fuel of the engine 1 may be any fuel as long as it is a liquid fuel containing at least gasoline.

(Engine configuration)
The engine 1 includes a cylinder block 12 and a cylinder head 13 that is placed on and fixed to the cylinder block 12. A plurality of cylinders 11 are formed inside the cylinder block 12. 1 and 3, only one cylinder 11 is shown. The engine 1 is a multi-cylinder engine.

  A piston 3 is slidably inserted in each cylinder 11. The piston 3 is connected to the crankshaft 15 via a connecting rod 14. The piston 3 defines a combustion chamber 17 together with the cylinder 11 and the cylinder head 13. The “combustion chamber” is not limited to the meaning of the space when the piston 3 reaches compression top dead center. The term “combustion chamber” may be used in a broad sense. That is, the “combustion chamber” may mean a space formed by the piston 3, the cylinder 11, and the cylinder head 13 regardless of the position of the piston 3.

  The upper surface of the piston 3 is a flat surface. A cavity 31 is formed on the upper surface of the piston 3. The cavity 31 is recessed from the upper surface of the piston 3. The cavity 31 faces an injector 6 described later.

  The cavity 31 has a convex portion 311. The convex portion 311 is provided at a position slightly displaced from the central axis X1 of the cylinder 11 toward the exhaust side. The convex part 311 is substantially conical. The convex portion 311 extends upward from the bottom of the cavity 31 along an axis X2 parallel to the central axis X1 of the cylinder 11. The upper end of the convex portion 311 is almost the same height as the upper surface of the peripheral edge of the cavity 31.

  The peripheral side surface of the cavity 31 is inclined with respect to the axis X <b> 2 from the bottom surface of the cavity 31 toward the opening of the cavity 31. The inner diameter of the cavity 31 gradually increases from the bottom of the cavity 31 toward the opening of the cavity 31.

  The cavity 31 has a bottom 313. A region on the intake side in the bottom 313 is opposed to a spark plug 25 described later. As shown in the upper diagram of FIG. 2, the bottom portion 313 extends in the horizontal direction so as to have a predetermined size in plan view.

  Further, the lower surface of the cylinder head 13, that is, the ceiling surface of the combustion chamber 17, is constituted by an intake-side inclined surface 1311 and an exhaust-side inclined surface 1312, as shown in the lower diagram of FIG. The inclined surface 1311 has an upward slope from the intake side toward the axis X2. On the other hand, the inclined surface 1312 has an upward slope from the exhaust side toward the axis X2. The ceiling surface of the combustion chamber 17 has a so-called pent roof shape.

  In addition, the shape of the combustion chamber 17 is not limited to the shape illustrated in FIG. For example, the shape of the cavity 31, the shape of the upper surface of the piston 3, the shape of the ceiling surface of the combustion chamber 17, and the like can be changed as appropriate.

  The cavity 31 may be symmetric with respect to the central axis X1 of the cylinder 11. The inclined surface 1311 and the inclined surface 1312 may be symmetric with respect to the central axis X1 of the cylinder 11.

  The geometric compression ratio of the engine 1 is set to 13 or more and 20 or less. As will be described later, the engine 1 performs an SPCCI combustion operation combining SI combustion and CI combustion in a part of the operation region. SPCCI combustion performs CI combustion operation using heat generation and pressure increase due to SI combustion. The engine 1 does not need to increase the temperature of the combustion chamber 17 (that is, the compression end temperature) when the piston 3 reaches the compression top dead center due to the self-ignition of the air-fuel mixture. That is, although the engine 1 performs the CI combustion operation, the value of the geometric compression ratio is set to be relatively low. Lowering the value of the geometric compression ratio is advantageous for reducing cooling loss and mechanical loss. The value of the geometric compression ratio of the engine 1 may be 14 to 17 in the regular specification (the fuel octane number is about 91), and may be 15 to 18 in the high-octane specification (the fuel octane number is about 96).

  An intake port 18 is formed in the cylinder head 13 for each cylinder 11. As shown in FIG. 3, the intake port 18 has two intake ports, a first intake port 181 and a second intake port 182. The first intake port 181 and the second intake port 182 are aligned in the axial direction of the crankshaft 15, that is, the front-rear direction of the engine 1. The intake port 18 communicates with the combustion chamber 17. Although not shown in detail, the intake port 18 is a so-called tumble port. That is, the intake port 18 has a shape in which a tumble flow is formed in the combustion chamber 17.

  An intake valve 21 is disposed in the intake port 18. The intake valve 21 opens and closes between the combustion chamber 17 and the intake port 18. The intake valve 21 is opened and closed at a predetermined timing by a valve operating mechanism. This valve mechanism can be a variable valve mechanism that makes the valve timing and / or valve lift variable. In this configuration example, as shown in FIGS. 1 and 4, the variable valve mechanism has an intake electric S-VT (Sequential-Valve Timing) 23. The intake motor S-VT 23 is configured to continuously change the rotation phase of the intake camshaft within a predetermined angle range. Thereby, the valve opening timing and the valve closing timing of the intake valve 21 change continuously. The valve mechanism of the intake valve 21 may have a hydraulic S-VT instead of the electric S-VT.

  The cylinder head 13 has an exhaust port 19 for each cylinder 11. The exhaust port 19 also has two exhaust ports, a first exhaust port 191 and a second exhaust port 192, as shown in FIG. The first exhaust port 191 and the second exhaust port 192 are arranged in the front-rear direction of the engine 1. The exhaust port 19 communicates with the combustion chamber 17. An exhaust valve 22 is disposed in the exhaust port 19. The exhaust valve 22 opens and closes between the combustion chamber 17 and the exhaust port 19. The exhaust valve 22 is opened and closed at a predetermined timing by a valve mechanism. This valve mechanism can be a variable valve mechanism that makes the valve timing and / or valve lift variable. In this configuration example, as shown in FIGS. 1 and 4, the variable valve mechanism has an exhaust electric S-VT 24. The exhaust electric S-VT 24 is configured to continuously change the rotational phase of the exhaust camshaft within a predetermined angle range. Thereby, the valve opening timing and the valve closing timing of the exhaust valve 22 change continuously. Note that the valve operating mechanism of the exhaust valve 22 may have a hydraulic S-VT instead of the electric S-VT.

  The engine 1 adjusts the length of the overlap period related to the opening timing of the intake valve 21 and the closing timing of the exhaust valve 22 by the intake electric S-VT 23 and the exhaust electric S-VT 24. Thereby, hot burned gas can be confined in the combustion chamber 17. That is, it becomes possible to introduce an internal EGR (Exhaust Gas Recirculation) gas into the combustion chamber 17. Further, the residual gas in the combustion chamber 17 can be scavenged by adjusting the length of the overlap period.

  An injector 6 is attached to the cylinder head 13 for each cylinder 11. The injector 6 is configured to inject fuel directly into the combustion chamber 17. The injector 6 is disposed facing the combustion chamber 17 in a valley portion of the pent roof where the intake-side inclined surface 1311 and the exhaust-side inclined surface 1312 intersect. As shown in FIG. 2, the injector 6 has an injection axis that is disposed in parallel to the center axis X <b> 1 of the cylinder. The injection axis of the injector 6 is coincident with the axis X2, and the injection axis of the injector 6 is coincident with the position of the convex portion 311 of the cavity 31. The injector 6 faces the cavity 31. The injection axis of the injector 6 may coincide with the central axis X1 of the cylinder 11. Also in this case, it is desirable that the injection axis of the injector 6 coincides with the position of the convex portion 311 of the cavity 31.

  Although not shown in detail, the injector 6 is constituted by a multi-injection type fuel injection valve having a plurality of injection holes. As shown by a plurality of regions consisting of two-dot chain lines in FIG. 2, the injector 6 has a fuel spray that spreads radially from the center of the combustion chamber 17 and spreads obliquely downward from the ceiling of the combustion chamber 17. Inject. In the present configuration example, the injector 6 has ten nozzle holes, and each nozzle hole is arranged at an equal angle in the circumferential direction. In addition, the number of nozzle holes may be eight. As shown in the upper diagram of FIG. 2, the nozzle hole is displaced in the circumferential direction with respect to a spark plug 25 described later. That is, the spark plug 25 is sandwiched between the shafts of two adjacent nozzle holes. Thereby, it is avoided that the spray of the fuel injected from the injector 6 directly hits the spark plug 25 and wets the electrode.

  A fuel supply system 61 is connected to the injector 6. The fuel supply system 61 includes a fuel tank 63 configured to store fuel, and a fuel supply path 62 that connects the fuel tank 63 and the injector 6 to each other. A fuel pump 65 and a common rail 64 are interposed in the fuel supply path 62. The fuel pump 65 pumps fuel to the common rail 64. In the present configuration example, the fuel pump 65 is, for example, an electric pump, and is disposed inside the fuel tank 63. The fuel pump 65 is connected to the fuel pump controller 651. The common rail 64 is configured to store the fuel pumped from the fuel pump 65 at a high fuel pressure. The common rail 64 is provided with a high pressure fuel pressure sensor SW16 and a fuel temperature sensor SW161. When the injector 6 is opened, the fuel stored in the common rail 64 is injected into the combustion chamber 17 from each injection port of the injector 6. The fuel supply system 61 is configured to be able to supply high pressure fuel of 30 MPa or more to the injector 6. The maximum fuel pressure of the fuel supply system 61 may be set to about 120 MPa, for example. The pressure of the fuel supplied to the injector 6 may be changed according to the operating state of the engine 1. The configuration of the fuel supply system 61 is not limited to the above configuration.

  A high-pressure fuel pump 641 and a low-pressure fuel pressure sensor SW20 are disposed between the common rail 64 and the fuel pump 65 in the fuel supply path 62 and upstream thereof. The high pressure fuel pump 641 is provided with a fuel temperature sensor SW21.

  A spark plug 25 is attached to the cylinder head 13 for each cylinder 11. The spark plug 25 forcibly ignites the air-fuel mixture in the combustion chamber 17. In the present configuration example, the spark plug 25 is disposed on the intake side across the central axis X1 of the cylinder 11 as shown in FIG. The spark plug 25 is adjacent to the injector 6. The spark plug 25 is located between the two intake ports 18. The spark plug 25 is attached to the cylinder head 13 so as to incline in a direction approaching the center of the combustion chamber 17 from above to below. The electrode of the spark plug 25 faces the combustion chamber 17 and is located near the ceiling surface of the combustion chamber 17.

  The cylinder head 13 is provided with a finger pressure sensor SW6 that detects the pressure in the combustion chamber 17 on the side opposite to the spark plug 25 (that is, the exhaust side) with respect to the center axis X1 of each cylinder 11.

  An intake passage 40 is connected to one side of the engine 1. The intake passage 40 communicates with the intake port 18 of each cylinder 11. The intake passage 40 is a passage through which the gas introduced into each combustion chamber 17 flows. An air cleaner 41 that filters fresh air is disposed at the upstream end of the intake passage 40. A surge tank (not shown) is disposed near the downstream end of the intake passage 40. The intake passage 40 downstream of the surge tank constitutes an independent passage that branches for each cylinder 11. The downstream end of the independent passage is connected to the intake port 18 of each cylinder 11.

  A throttle valve 43 is disposed between the air cleaner 41 and the surge tank in the intake passage 40. The throttle valve 43 is configured to adjust the amount of fresh air introduced into the combustion chamber 17 by adjusting the opening of the valve.

  In the intake passage 40, a supercharger 44 is disposed downstream of the throttle valve 43. The supercharger 44 is configured to supercharge the gas introduced into the combustion chamber 17. In the present configuration example, the supercharger 44 is a mechanical supercharger that is driven by the engine 1. The mechanical supercharger 44 may be, for example, a roots type. The configuration of the mechanical supercharger 44 may be any configuration. The mechanical supercharger 44 may be, for example, a Rishorum type, a vane type, or a centrifugal type.

  An electromagnetic clutch 45 is interposed between the supercharger 44 and the engine 1. The electromagnetic clutch 45 transmits a driving force from the engine 1 to the supercharger 44 between the supercharger 44 and the engine 1 or interrupts the transmission of the driving force. As will be described later, when the ECU 10 switches between disconnection and connection of the electromagnetic clutch 45, the supercharger 44 is switched on and off. The engine 1 is configured to be able to switch between supercharging the gas introduced into the combustion chamber 17 by the supercharger 44 and not supercharging the gas introduced into the combustion chamber 17 by the supercharger 44. ing.

  An intercooler 46 is disposed downstream of the supercharger 44 in the intake passage 40. The intercooler 46 is configured to cool the gas compressed in the supercharger 44. The intercooler 46 may be configured to be, for example, a water cooling type.

  A bypass passage 47 is connected to the intake passage 40. The bypass passage 47 connects the upstream portion of the supercharger 44 and the downstream portion of the intercooler 46 in the intake passage 40 so as to bypass the supercharger 44 and the intercooler 46. An air bypass valve 48 is disposed in the bypass passage 47. The air bypass valve 48 adjusts the flow rate of the gas flowing through the bypass passage 47.

  When the supercharger 44 is turned off, that is, when the electromagnetic clutch 45 is disconnected, the air bypass valve 48 is fully opened. As a result, the gas flowing through the intake passage 40 bypasses the supercharger 44 and is introduced into the combustion chamber 17 of the engine 1. In this case, the engine 1 is operated in a non-supercharged state, that is, in a natural intake state.

  When the supercharger 44 is turned on, that is, when the electromagnetic clutch 45 is connected, a part of the gas that has passed through the supercharger 44 flows back upstream of the supercharger 44 through the bypass passage 47. This reverse flow rate can be adjusted by adjusting the opening degree of the air bypass valve 48, so that the supercharging pressure of the gas introduced into the combustion chamber 17 can be adjusted. In the present configuration example, the supercharger 44, the bypass passage 47 and the air bypass valve 48 constitute a supercharging system 49.

  The engine 1 has a swirl generator that generates a swirl flow in the combustion chamber 17. An example of the swirl generator is a swirl control valve 56 attached to the intake passage 40 as shown in FIG. The swirl control valve 56 is disposed in the secondary passage 402 among the primary passage 401 connected to the first intake port 181 and the secondary passage 402 connected to the second intake port 182. The swirl control valve 56 is an opening adjustment valve that can narrow the cross section of the secondary passage 402. When the opening of the swirl control valve 56 is small, the intake flow rate flowing into the combustion chamber 17 from the first intake port 181 among the first intake port 181 and the second intake port 182 aligned in the front-rear direction of the engine 1 is On the other hand, the flow rate of intake air flowing into the combustion chamber 17 from the second intake port 182 relatively decreases, so that the swirl flow in the combustion chamber 17 becomes stronger. When the opening of the swirl control valve 56 is large, the intake flow rate flowing into the combustion chamber 17 from each of the first intake port 181 and the second intake port 182 becomes substantially equal, so that the swirl flow in the combustion chamber 17 is become weak. When the swirl control valve 56 is fully opened, no swirl flow is generated. The swirl flow circulates counterclockwise in FIG. 3 as indicated by the arrows (see also the white arrows in FIG. 2).

  The swirl generating unit shifts the valve opening periods of the two intake valves 21 instead of attaching the swirl control valve 56 to the intake passage 40 or in addition to attaching the swirl control valve 56, and only one of the intake valves 21. Alternatively, a configuration in which intake air can be introduced into the combustion chamber 17 may be employed. Since only one of the two intake valves 21 is opened, intake air is unevenly introduced into the combustion chamber 17, so that a swirl flow can be generated in the combustion chamber 17. it can. Further, the swirl generator may be configured to generate a swirl flow in the combustion chamber 17 by devising the shape of the intake port 18.

  An exhaust passage 50 is connected to the other side of the engine 1. The exhaust passage 50 communicates with the exhaust port 19 of each cylinder 11. The exhaust passage 50 is a passage through which the exhaust discharged from the combustion chamber 17 flows. Although the detailed illustration is omitted, the upstream portion of the exhaust passage 50 constitutes an independent passage branched for each cylinder 11. The upstream end of the independent passage is connected to the exhaust port 19 of each cylinder 11.

  The exhaust passage 50 is provided with an exhaust purification system having a plurality of catalytic converters. Although not shown, the upstream catalytic converter is disposed in the engine room. The upstream catalytic converter includes a three-way catalyst 511 and a GPF (Gasoline Particulate Filter) 512. The downstream catalytic converter is disposed outside the engine room. The downstream catalytic converter has a three-way catalyst 513. The exhaust purification system is not limited to the configuration shown in the figure.

  An EGR passage 52 constituting an external EGR system is connected between the intake passage 40 and the exhaust passage 50. The EGR passage 52 is a passage for returning a part of the burned gas to the intake passage 40. The upstream end of the EGR passage 52 is connected between the upstream catalytic converter and the downstream catalytic converter in the exhaust passage 50. The downstream end of the EGR passage 52 is connected to the upstream side of the supercharger 44 in the intake passage 40.

  A water-cooled EGR cooler 53 is disposed in the EGR passage 52. The EGR cooler 53 is configured to cool the burned gas. An EGR valve 54 is also disposed in the EGR passage 52. The EGR valve 54 is configured to adjust the flow rate of burned gas flowing through the EGR passage 52. By adjusting the opening degree of the EGR valve 54, the recirculation amount of the cooled burned gas, that is, the external EGR gas can be adjusted.

  In this configuration example, the EGR system 55 includes an external EGR system configured to include the EGR passage 52 and the EGR valve 54, and an internal configuration configured to include the intake electric S-VT 23 and the exhaust electric S-VT 24 described above. And an EGR system.

  The control device for the compression self-ignition engine includes an ECU (Engine Control Unit) 10 for operating the engine 1. The ECU 10 is a controller based on a well-known microcomputer, and as shown in FIG. 4, a central processing unit (CPU) 101 that executes a program, and, for example, a RAM (Random Access Memory) and a ROM The memory 102 is configured by (Read Only Memory) and stores a program and data, and an input / output bus 103 that inputs and outputs electric signals. The ECU 10 is an example of a control unit.

  As shown in FIGS. 1 and 4, various sensors such as SW1 to SW17, SW20 to SW24, SW31, SW51, SW101, SW102, and SW161 are connected to the ECU 10. Each of the above sensors outputs a detection signal to the ECU 10. The sensor includes, for example, the following plurality of sensors.

That is, the air flow sensor SW1 that is disposed downstream of the air cleaner 41 in the intake passage 40 and detects the flow rate of fresh air flowing through the intake passage 40, the first intake temperature sensor SW2 that detects the temperature of fresh air, and the EGR in the intake passage 40 An intake pressure sensor SW3 that is disposed downstream of the connection position of the passage 52 and upstream of the supercharger 44, detects the pressure of the gas flowing into the supercharger 44, and a second temperature of the gas. An intake air temperature sensor SW31, a third intake air temperature sensor SW4 that is disposed downstream of the supercharger 44 in the intake passage 40 and upstream of the intercooler 46, and detects the temperature of the gas flowing out of the supercharger 44; A boost pressure sensor SW5 that is attached to a surge tank downstream of the cooler 46 and detects the pressure of the gas downstream of the supercharger 44, and the gas A fourth intake air temperature sensor SW51 for detecting the degree of pressure, and a finger pressure sensor SW6 for detecting the pressure in each combustion chamber 17, respectively attached to the cylinder head 13 corresponding to each cylinder 11, and disposed in the exhaust passage 50. An exhaust gas temperature sensor SW7 for detecting the temperature of exhaust gas discharged from the exhaust gas, a linear O ₂ sensor SW8 for detecting the oxygen concentration in the exhaust gas, upstream of the upstream catalytic converter in the exhaust passage 50, and three in the upstream catalytic converter. A lambda O ₂ sensor SW9 that is disposed downstream of the original catalyst 511 and detects the oxygen concentration in the exhaust gas after passing through the three-way catalyst 511, and a first water temperature sensor SW10 that is attached to the cylinder head 13 and detects the temperature of the cooling water. (Refer to the upper right figure), attached near the outlet of the cylinder head 13 toward the main radiator, A second water temperature sensor SW101 for detecting the temperature, a third water temperature sensor SW102 for detecting the temperature of the cooling water, attached to the vicinity of the outlet toward the water pump W / P of the main radiator, and attached to the engine 1, A crank angle sensor SW11 for detecting a rotation angle, attached to an accelerator pedal mechanism, an accelerator opening sensor SW12 for detecting an accelerator opening corresponding to the amount of operation of the accelerator pedal, an engine 1 and an intake camshaft rotation angle An intake cam angle sensor SW13 that detects the rotational angle of the exhaust camshaft, is attached to the engine 1 and is disposed in the EGR passage 52, and detects a differential pressure upstream and downstream of the EGR valve 54. EGR differential pressure sensor SW15, common rail of fuel supply system 61 4, a fuel pressure sensor SW 16 for detecting the pressure of fuel supplied to the injector 6, a fuel temperature sensor SW 161 for detecting the temperature of the fuel, and a drive motor for the throttle valve 43. Is a throttle opening sensor SW17.

  Further, the low pressure fuel pressure sensor SW20 attached between the high pressure fuel pump 641 and the fuel pump 65 in the fuel supply path 62, the fuel temperature sensor SW21 attached to the high pressure fuel pump 641, and the GPF pressure sensor SW23 attached to the GPF 512. And a hydraulic pressure sensor SW23 attached to the cylinder block 22 and an oil level sensor SW24 attached to the bottom of the oil pan.

  The ECU 10 determines the operating state of the engine 1 based on these detection signals and calculates the control amount of each device. The ECU 100 sends control signals relating to the calculated control amount to the injector 6, spark plug 25, intake electric S-VT 23, exhaust electric S-VT 24, fuel supply system 61, throttle valve 43, EGR valve 54, supercharger 44. Are output to the electromagnetic clutch 45, the air bypass valve 48, and the swirl control valve 56, respectively. For example, the ECU 10 adjusts the supercharging pressure by adjusting the opening degree of the air bypass valve 48 based on the differential pressure across the supercharger 44 obtained from the detection signals of the intake pressure sensor SW3 and the boost pressure sensor SW5. . Further, the ECU 10 adjusts the opening degree of the EGR valve 54 based on the differential pressure across the EGR valve 54 obtained from the detection signal of the EGR differential pressure sensor SW15, whereby the amount of external EGR gas introduced into the combustion chamber 17 is adjusted. Adjust. Details of control of the engine 1 by the ECU 10 will be described later.

(Engine operating range)
FIG. 5 exemplifies operation region maps 501 and 502 of the engine 1. The operation region maps 501 and 502 of the engine 1 are determined by the load and the rotational speed, and are divided into five regions for the load level and the rotational speed level. Specifically, these five regions include idle operation and have a low load region (1) -1 that extends to a region of low rotation and medium rotation, a load higher than the low load region, and low rotation and Medium-load region (1) -2 spreading in the medium-rotation region, medium-load region (2) in the high-load region that is a region where the load is higher than the medium-load region (1) -2 and includes a fully open load, high A low rotation region (3) having a lower rotational speed than the medium rotation region (2) in the load region, a low load region (1) -1, a medium load region (1) -2, a high load medium rotation region (2), And a high rotation region (4) having a higher rotational speed than the high load low rotation region (3). Here, the low rotation region, the medium rotation region, and the high rotation region are obtained when the entire operation region of the engine 1 is divided into approximately three equal parts of the low rotation region, the medium rotation region, and the high rotation region, respectively. A low rotation region, a medium rotation region, and a high rotation region may be used. In the example of FIG. 5, the rotation speed less than N1 is low rotation, the rotation speed N2 or more is high rotation, and the rotation speed N1 or more and less than N2 is medium rotation. For example, the rotational speed N1 may be about 1200 rpm, and the rotational speed N2 may be about 4000 rpm, for example. Further, the high load mid-rotation region (2) may be a region where the combustion pressure is 900 kPa or more. In FIG. 5, for easy understanding, the operation region maps 501 and 502 of the engine 1 are drawn in two parts. A map 501 shows the state of the air-fuel mixture and the combustion mode in each region, and the drive region and non-drive region of the supercharger 44. A map 502 shows the opening degree of the swirl control valve 56 in each region. Note that a two-dot chain line in FIG. 5 indicates a road-load line of the engine 1.

  The engine 1 performs compression self-ignition in a low load region (1) -1, a medium load region (1) -2, and a high load medium rotation region (2) mainly for the purpose of improving fuel consumption and exhaust gas performance. The combustion operation (SPCCI combustion) is performed.

  In the SPCCI combustion, the spark plug 25 forcibly ignites the air-fuel mixture in the combustion chamber 17 so that the air-fuel mixture undergoes SI combustion by flame propagation, and the heat in the combustion chamber 17 generates heat from the SI combustion. This is a form in which the unburned mixture undergoes CI combustion due to self-ignition as the temperature rises and the pressure in the combustion chamber 17 increases due to flame propagation.

  By adjusting the calorific value of the SI combustion, the temperature variation in the combustion chamber 17 before the start of compression can be absorbed. Even if the temperature in the combustion chamber 17 before the compression starts varies, for example, if the SI combustion start timing is adjusted by adjusting the ignition timing, the air-fuel mixture can be self-ignited at the target timing.

  In SPCCI combustion, heat generation during SI combustion is milder than heat generation during CI combustion. In the waveform of the heat generation rate in the SPCCI combustion, as shown by reference numerals 6014, 6024, 6034, and 6063 in FIG. The pressure fluctuation (dp / dθ) in the combustion chamber 17 is also gentler during SI combustion than during CI combustion. In other words, the heat generation rate waveform of SPCCI combustion includes the first heat generation rate portion formed by SI combustion with a relatively small rising inclination and the second heat generation rate portion formed by CI combustion with a relatively large rising inclination. The heat generation rate part is formed to be continuous in this order.

  When the temperature and pressure in the combustion chamber 17 are increased by SI combustion, the unburned mixture is self-ignited. In the examples of the heat generation rate waveforms 6014, 6024, 6034, and 6063 shown in FIG. 6, the slope of the waveform changes from small to large at the timing of self-ignition. That is, the heat generation rate waveform has an inflection point at the timing when CI combustion starts.

  After the start of CI combustion, SI combustion and CI combustion are performed in parallel. Since CI combustion generates more heat than SI combustion, the heat generation rate is relatively large. However, since CI combustion is performed after compression top dead center, the piston 3 is lowered by motoring. It is avoided that the slopes of the heat generation rate waveforms 6014, 6024, 6034, and 6063 due to CI combustion become too large. Dp / dθ at the time of CI combustion also becomes relatively gentle.

  dp / dθ can be used as an index representing combustion noise. However, since SPCCI combustion can reduce dp / dθ as described above, it is possible to avoid excessive combustion noise. . Combustion noise can be suppressed below an acceptable level.

  When CI combustion ends, SPCCI combustion ends. CI combustion has a shorter combustion period than SI combustion. In SPCCI combustion, the combustion end timing is earlier than SI combustion. In other words, SPCCI combustion can bring the combustion end time during the expansion stroke closer to the compression top dead center. The SPCCI combustion is more advantageous for improving the fuel efficiency of the engine 1 than the SI combustion.

  Further, in SPCCI combustion, the start timing of CI combustion (second heat generation rate part) is determined by changing the heat generation amount of SI combustion (first heat generation rate part) according to the operating state of the engine. Combustion control means (EGR, VVT, intake air amount control means) are controlled so that the target CI combustion start timing set according to the state is reached.

  The engine 1 also performs SI combustion operation by spark ignition in other regions, that is, the high load low rotation region (3) and the high rotation region (4). Hereinafter, the operation of the engine 1 in each region will be described in detail with reference to the fuel injection timing and the ignition timing shown in FIG.

  FIG. 6 shows the combustion injection timing and ignition timing in each region of the operation region maps 501 and 502 of FIG. Reference numerals 601, 602, 603, 604, 605, and 606 in FIG. 6 correspond to the operating states 601, 602, 602, 603, 604, 605, and 606 in FIG. 5, respectively. The operation state 606 corresponds to an operation state in which the number of rotations is high in the high load mid-rotation region (2).

(Low load area (1) -1)
When the engine 1 is operating in the low load region (1) -1, the engine 1 performs the CI combustion operation as described above. In the combustion by self-ignition, when the temperature in the combustion chamber 17 before the start of compression varies, the timing of self-ignition greatly changes. Therefore, the engine 1 performs an SPCCI combustion operation that combines SI combustion and CI combustion in the low load region (1) -1.

  When the engine 1 is operated in the operation state 601 in the low load region (1) -1, the injector 6 divides the fuel into a plurality of times during the compression stroke and injects the fuel into the combustion chamber 17 (reference numeral 6015, 6016). Due to the split fuel injection and the strong swirl flow in the combustion chamber 17, the air-fuel mixture is stratified in the central portion and the outer peripheral portion of the combustion chamber 17.

  After completion of fuel injection, the spark plug 25 ignites the air-fuel mixture at the center of the combustion chamber 17 at a predetermined timing before compression top dead center (see reference numeral 6013). Since the air-fuel mixture in the central portion has a relatively high fuel concentration, the ignitability is improved and SI combustion by flame propagation is stabilized. By stabilizing the SI combustion, the CI combustion starts at an appropriate timing (see the combustion waveform 6014). In SPCCI combustion, controllability of CI combustion is improved. As a result, when the engine 1 operates in the low load region (1) -1, it is possible to achieve both the suppression of the generation of combustion noise and the improvement of fuel consumption performance by shortening the combustion period.

  In the low load region (1) -1, since the engine 1 performs the SPCCI combustion operation with the air-fuel mixture leaner than the stoichiometric air-fuel ratio, the low load region (1) -1 is "SPCCIλ> 1 region". Can be called. That is, the “second combustion mode”.

(Medium load area (1) -2)
Even when the engine 1 is operating in the medium load region (1) -2, the engine 1 performs the SPCCI combustion operation similarly to the low load region (1) -1.

  When the engine 1 operates in the operation state 602 in the medium load region (1) -2, the injector 6 performs fuel injection during the intake stroke (see reference numeral 6025) and fuel injection during the compression stroke (reference numeral 6026). See). By performing the first injection 6025 during the intake stroke, the fuel can be distributed substantially uniformly in the combustion chamber 17. By performing the second injection 6026 during the compression stroke, when the load is high in the middle load region (1) -2, the temperature in the combustion chamber 17 is lowered by the latent heat of vaporization of the fuel to cause abnormal combustion such as knocking. Can be prevented. For example, the ratio of the injection amount of the first injection 6025 and the injection amount of the second injection 6026 may be 95: 5.

  When the injector 6 performs the first injection 6025 during the intake stroke and the second injection 6026 during the compression stroke, the excess air ratio λ is 1.0 ± 0.2 as a whole in the combustion chamber 17. A mixed gas mixture is formed. Since the fuel concentration of the air-fuel mixture is substantially uniform, it is possible to improve fuel efficiency by reducing unburned loss and to improve exhaust gas performance by avoiding the generation of smoke. The excess air ratio λ is preferably 1.0 to 1.2.

  When the spark plug 25 ignites the air-fuel mixture at a predetermined timing before the compression top dead center (see reference numeral 6023), the air-fuel mixture burns by flame propagation. After combustion by flame propagation starts, the unburned mixture self-ignites at the target timing and performs CI combustion (see combustion waveform 6024).

  In the middle load region (1) -2, the engine 1 performs the SPCCI combustion operation with the air-fuel mixture as the stoichiometric air-fuel ratio, and therefore, the middle load region (1) -2 can be referred to as “SPCCIλ = 1 region”. . That is, the “first combustion mode”.

  Here, as shown in a map 501 in FIG. 5, the supercharger 44 is turned off in a part of the low load region (1) -1 and a part of the medium load region (1) -2 ( (Refer to the description of “S / C OFF” in the figure.) Specifically, the supercharger 44 is turned off in the low rotation side region in the low load region (1) -1. In the region on the high rotation side in the low load region (1) -1, the supercharger 44 is turned on in order to ensure a necessary intake charge amount corresponding to the increase in the rotational speed of the engine 1. Increase the boost pressure. Further, the supercharger 44 is turned off in the low load / low rotation side region in the medium load region (1) -2, and the fuel injection amount in the high load side region in the medium load region (1) -2. The turbocharger 44 is turned on in order to ensure a necessary intake charge amount corresponding to the increase in the amount of air. Also, in the region on the high speed side, the supercharger 44 is turned on in order to ensure a necessary intake charge amount corresponding to the increase in the rotational speed of the engine 1.

  Note that, in each of the high load mid-rotation region (2), the high load low rotation region (3), and the high rotation region (4), the supercharger 44 is turned on over the entire region.

(High load medium rotation range (2))
Even when the engine 1 is operating in the high load mid-rotation region (2), the engine 1 performs the SPCCI combustion operation similarly to the low load region (1) -1 and the medium load region (1) -2. .

  When the engine 1 is operated in the operation state 603 on the low rotation side in the high load mid-rotation region (2), the injector 6 injects fuel in the intake stroke (see reference numeral 6035) and at the end of the compression stroke. Fuel is injected (see reference numeral 6036).

  The pre-injection 6035 that starts in the intake stroke may start fuel injection in the first half of the intake stroke. The first half of the intake stroke may be the first half when the intake stroke is divided into two equal parts, the first half and the second half. Specifically, the front injection may start fuel injection at 280 ° CA before top dead center.

  When the injection start of the front injection 6035 is performed in the first half of the intake stroke, the fuel spray strikes the opening edge of the cavity 31, so that a part of the fuel enters the squish area 171 of the combustion chamber 17, and the remaining fuel is in the cavity 31. to go into. The swirl flow is strong at the outer periphery of the combustion chamber 17 and weak at the center. Therefore, a part of the fuel that has entered the squish area 171 enters the swirl flow, and the remaining fuel that has entered the cavity 31 enters the inside of the swirl flow. The fuel that has entered the swirl flow remains in the swirl flow during the intake stroke to the compression stroke, and forms an air-fuel mixture for CI combustion at the outer peripheral portion of the combustion chamber 17. The fuel that has entered the swirl flow also remains inside the swirl flow during the intake stroke to the compression stroke, and forms an air-fuel mixture for SI combustion in the central portion of the combustion chamber 17.

  The air-fuel ratio in the central portion where the spark plug 25 is disposed preferably has an excess air ratio λ of 1 or less, and the air-fuel mixture in the outer peripheral portion has an air excess ratio λ of 1 or less, preferably less than 1. The air-fuel ratio (A / F) of the air-fuel mixture in the center may be, for example, 13 or more and the theoretical air-fuel ratio (14.7) or less. The air-fuel ratio of the air-fuel mixture at the center may be leaner than the stoichiometric air-fuel ratio. The air-fuel ratio of the air-fuel mixture at the outer peripheral portion may be, for example, 11 or more and the stoichiometric air-fuel ratio or less, preferably 11 or more and 12 or less. The air-fuel ratio of the entire air-fuel mixture in the combustion chamber 17 may be 12.5 or more and the stoichiometric air-fuel ratio or less, preferably 12.5 or more and 13 or less.

  The post-injection 6036 performed at the end of the compression stroke may start the fuel injection at 10 ° CA before top dead center, for example. By performing the post-stage injection immediately before the top dead center, the temperature in the combustion chamber can be lowered by the latent heat of vaporization of the fuel. The fuel injected by the upstream injection 6035 undergoes a low-temperature oxidation reaction during the compression stroke, and shifts to a high-temperature oxidation reaction before top dead center. At this time, by performing the post-injection 6036 immediately before the top dead center and lowering the temperature in the combustion chamber, it is possible to suppress the transition from the low temperature oxidation reaction to the high temperature oxidation reaction, thereby suppressing the occurrence of premature ignition. can do. In addition, the ratio of the injection quantity of the front | former stage injection 6035 and the injection quantity of the back | latter stage injection 6036 is good also as 95: 5 as an example.

  The spark plug 25 ignites the air-fuel mixture in the center of the combustion chamber 17 in the vicinity of the compression top dead center (see reference numeral 6037). The spark plug 25 ignites, for example, after compression top dead center. Since the spark plug 25 is disposed at the center of the combustion chamber 17, the air-fuel mixture at the center starts SI combustion by flame propagation by the ignition of the spark plug 25. The flame of SI combustion propagates in the circumferential direction on a strong swirl flow in the combustion chamber 17. The unburned air-fuel mixture undergoes compression ignition at a predetermined circumferential position in the outer peripheral portion of the combustion chamber 17, and CI combustion starts (see combustion waveform 6034).

  On the other hand, when the engine 1 is operated in the high-rotation-side operation state 606 in the high load mid-rotation region (2), the injector 6 starts fuel injection in the intake stroke (see reference numeral 6061).

  The front injection 6061 that starts in the intake stroke may start fuel injection in the first half of the intake stroke in the same manner as the front injection 6035 in the operating state 603. Specifically, the front injection 6061 may start fuel injection at 280 ° CA before top dead center. The end of the pre-injection 6061 may be in the compression stroke beyond the intake stroke. By making the injection start of the front injection 6061 in the first half of the intake stroke, a mixture for CI combustion is formed in the outer peripheral portion of the combustion chamber 17 and a mixture for SI combustion is formed in the central portion of the combustion chamber 17. be able to. Since the rotational speed is high, it is difficult for abnormal combustion to occur, so that the post-injection can be omitted.

  The spark plug 25 ignites the air-fuel mixture at the center of the combustion chamber 17 in the vicinity of the compression top dead center (see reference numeral 6062). The spark plug 25 ignites, for example, after compression top dead center. Thereby, SPCCI combustion is performed (see combustion waveform 6063).

  In the high load mid-rotation region (2), the engine 1 performs the SPCCI combustion operation in a state where the air-fuel mixture is richer than the stoichiometric air-fuel ratio. Therefore, the high load mid-rotation region (2) is “SPCCIλ ≦ 1 region”. Can be called.

(High load, low rotation range (3))
When the rotational speed of the engine 1 is low, the time required for the crank angle to change by 1 ° becomes long. In the high load low rotation region (3), as in the high load medium rotation region (2), for example, when fuel is injected into the combustion chamber 17 in the first half of the intake stroke and the compression stroke, the reaction of the fuel proceeds excessively. This may lead to premature ignition. When the engine 1 is operating in the high-load low-rotation region (3), it is difficult to perform the aforementioned SPCCI combustion operation.

  Therefore, when the engine 1 is operating in the high-load low-rotation region (3), the engine 1 performs the SI combustion operation instead of the SPCCI combustion operation.

  When the engine 1 is operated in the operation state 604 in the high-load low-rotation region (3), the injector 6 is in the combustion chamber at each timing of the intake stroke and the retard period from the end of the compression stroke to the beginning of the expansion stroke. 17 is injected into the fuel (see reference numerals 6044 and 6045). In this way, by injecting fuel in two steps, the amount of fuel injected within the retard period can be reduced. By injecting fuel during the intake stroke (see reference numeral 6044), it is possible to secure a sufficient time for forming the air-fuel mixture. Also, by injecting fuel during the retard period (see reference numeral 6045), the flow in the combustion chamber 17 can be increased immediately before ignition, which is advantageous for stabilizing SI combustion. This form of fuel injection is particularly effective when the value of the geometric compression ratio of the engine 1 is low.

  After the fuel is injected, the spark plug 25 ignites the air-fuel mixture at a timing near the compression top dead center (see reference numeral 6042). For example, the spark plug 25 may perform ignition after compression top dead center. The air-fuel mixture undergoes SI combustion in the expansion stroke. Since SI combustion starts in the expansion stroke, CI combustion does not start (see combustion waveform 6043).

  In the high-load low-rotation region (3), the engine 1 performs SI combustion operation by injecting fuel during the retard period from the end of the compression stroke to the early stage of the expansion stroke. It can be called “SI region”.

(High rotation area (4))
When the rotation speed of the engine 1 is high, the time required for the crank angle to change by 1 ° is shortened. For this reason, for example, it becomes difficult to stratify the air-fuel mixture in the combustion chamber 17 by performing split injection during the compression stroke in the high rotation region in the high load region. That is, when the rotational speed of the engine 1 is increased, it becomes difficult to perform the aforementioned SPCCI combustion operation. For this reason, when the engine 1 is operating in the high speed region (4), the engine 1 performs the SI combustion operation instead of the SPCCI combustion operation. The high rotation region (4) extends over the entire load direction from a low load to a high load.

  Reference numeral 605 in FIG. 6 denotes a fuel injection timing (see reference numeral 6051) and an ignition timing (see reference numeral 6052) and combustion when the engine 1 is operating in a high load region (4). An example of each of the waveforms (see reference numeral 6053) is shown.

  The EGR system 55 introduces EGR gas into the combustion chamber 17 when the operating state of the engine 1 is in the high rotation region (4). The engine 1 reduces the amount of EGR gas as the load increases. At full load, EGR gas may be zero.

  When the engine 1 is operated in the high speed region (4), the swirl control valve (SCV) 56 is fully opened. A swirl flow is not generated in the combustion chamber 17 and only a tumble flow is generated. By fully opening the swirl control valve 56, the charging efficiency can be increased in the high rotation region (4), and the pump loss can be reduced.

  When the engine 1 operates in the high speed region (4), the value of the air-fuel ratio (A / F) of the air-fuel mixture is basically the stoichiometric air-fuel ratio (A / F = 14. 7). That is, the excess air ratio λ of the air-fuel mixture may be 1.0 ± 0.2. Note that the excess air ratio λ of the air-fuel mixture may be less than 1 in the high load region including the fully open load in the high rotation region (4).

  When the engine 1 operates in the high speed region (4), the injector 6 starts fuel injection during the intake stroke. The injector 6 injects fuel in a lump. 6 exemplifies the fuel injection state when the load of the engine 1 is high and the fuel injection amount is large, and the fuel injection period changes according to the fuel injection amount. By starting fuel injection during the intake stroke, a homogeneous or substantially homogeneous air-fuel mixture can be formed in the combustion chamber 17. Moreover, since the fuel vaporization time can be ensured as long as possible when the engine 1 is rotating at a high speed, unburned loss can be reduced and soot generation can be suppressed.

  The spark plug 25 ignites the air-fuel mixture at an appropriate timing before the compression top dead center after the fuel injection ends (see reference numeral 6052).

  In the high speed region (4), the engine 1 starts the fuel injection during the intake stroke and performs the SI combustion operation. Therefore, the high speed region (4) can be referred to as an “intake-SI region”.

("Reduce cylinder operation" in low load range (1) -1)
In the present embodiment, as shown in FIG. 5, the reduced-cylinder operation is performed in the lowest low load region of the low load region (1) -1 (referred to here as “layer 3” for convenience). The lowest low load region of layer 3 may be, for example, a region of 200 kPa or less in terms of axial average effective pressure (BMEP). In addition, a shaft mean effective pressure (BMEP: Break Mean Effective Pressure) does not represent the load itself, but a value obtained by multiplying the BMEP by the displacement is proportional to the shaft torque.

  In this way, in an operating region where BMEP is 200 kPa or less, for example, throttle loss (pump loss) during combustion of the engine 1 becomes large. Reduced cylinder operation (cylinder deactivation operation) is performed. At this time, the throttle valve 43 of the inactive cylinder is opened. For the cylinders in operation, the value of the mixture compression ratio (geometric compression ratio) may be set to a relatively high value of 13.0 or more as described above.

  By such a reduced cylinder operation, the load on each of the other two cylinders in operation increases, so that throttle loss can be suppressed.

  In addition, it is good also as a state which act | operated the intake valve and the exhaust valve with respect to the cylinder in a pause. In this way, it is not necessary to provide a valve stop mechanism, so that it is possible to reduce the manufacturing cost of an engine capable of reducing cylinder operation.

  During the reduced-cylinder operation, supercharging by the supercharger 44 is not performed, but the electromagnetic clutch 45 of the supercharger 44 may be kept connected. This is because the mechanical load due to the connection and disconnection of the electromagnetic clutch 45 becomes smaller when the electromagnetic clutch 45 is connected regardless of the rotational speed of the engine 1.

  Further, as shown in a map 502 in FIG. 5, during the reduced-cylinder operation, a swirl flow may be generated for the cylinder in operation. In this case, the value of the swirl ratio in the swirl flow may be set to 4 or more. This makes it possible to stably perform SI combustion with an excess air ratio λ> 1, that is, A / F lean, and to enhance the swirl flow generated in the combustion chamber, further stabilizing SI combustion. can do.

  When the intake air temperature to the combustion chamber is lower than a predetermined value, the all-cylinder operation may be performed instead of the reduced-cylinder operation. This eliminates the instability of SI combustion when the intake air temperature is low, so that it is possible to suppress a delay in the combustion timing of subsequent CI combustion.

  Further, when the coolant temperature of the engine 1 is lower than a predetermined value, an all-cylinder operation may be performed instead of the reduced-cylinder operation. As a result, in the all-cylinder operation, the cooling of the end gas (unburned mixture) in the wall surface of the combustion chamber and in the vicinity thereof is quickly eliminated as compared with the case of the reduced-cylinder operation. Delay can be suppressed.

  Further, as shown in FIG. 5, in the medium load region (1) -2 (herein referred to as “layer 2” for the sake of convenience), an air-fuel ratio equivalent to the value of the theoretical air-fuel ratio (that is, λ = 1). Thus, SI combustion and CI combustion following the SI combustion are performed.

  Here, even when transition is made from layer 2 performed at an air-fuel ratio equivalent to the excess air ratio λ = 1 to layer 3 depending on the operating state of the engine 1, the reduced-cylinder operation may not be performed. That is, even if the engine 1 is operating at the layer 3 with λ = 1, even if the engine 1 shifts to the layer 3, it does not shift to the reduced-cylinder operation and continues the all-cylinder operation as it is. Thereby, generation | occurrence | production of NOx at the time of transfer of control to reduced cylinder operation can be suppressed.

  Next, the relationship between the load and the fuel amount in layer 2 and layer 3 will be described with reference to the drawings.

  FIGS. 7 to 9 show an example of the relationship between the first injection timing, the second injection timing, the third injection timing and each load in the layer 3 including the reduced cylinder operation (cylinder deactivation) and the layer 2. . 10 to 12 show an example of the relationship between the amount of fuel per cycle and the load for each of a plurality of fuel injection timings. Here, the rotation speed of the engine 1 in each graph of FIGS. 7 to 12 is, for example, about 3500 rpm. The same applies to FIGS. 13 to 16 described later.

  As shown in FIG. 7, for example, of the three injection timings, the first injection timing is about 88 ° CA before compression top dead center (bTDC) in the low load region where all cylinder operation of layer 3 is performed. Also good. At the load time point when shifting to the layer 3 where the cylinder is rested at a lower load, the first injection timing may be about 320 ° CA before the compression top dead center, and is advanced to the initial stage of the intake stroke. As the load decreases, the first injection timing may be retarded to about 90 ° CA before compression top dead center.

  On the other hand, as shown in FIG. 10, the first injected fuel amount gradually decreases as the load decreases in the layer 3 that performs the all-cylinder operation. Doubled. As the load further decreases, the first injected fuel amount decreases.

  As shown in FIG. 8, the second injection timing among the three injection timings is about 60 ° CA to about 50 ° CA before compression top dead center in the low load region in which all cylinder operation of layer 3 is performed. May be. At the load time point when shifting to the layer 3 where the cylinder is rested at a lower load, the second injection timing may be about 250 ° CA before compression top dead center. As the load decreases, the second injection timing may be retarded to about 60 ° CA before compression top dead center.

  On the other hand, as shown in FIG. 11, the second injected fuel amount gradually decreases in the low load region where the all-cylinder operation of the layer 3 is performed, and further decreases by one step at the load time point when shifting to cylinder deactivation. . As the load further decreases, the amount of the second injected fuel is increased.

  As shown in FIG. 9, the third injection timing among the three injection timings may be about 45 ° CA before compression top dead center in the low load region where the all-cylinder operation of the layer 3 is performed. At the load time point when shifting to the layer 3 where the cylinder load is lowered, the third injection timing may be approximately 200 ° CA before compression top dead center. As the load decreases, the third injection timing may be retarded to about 55 ° CA before compression top dead center.

  On the other hand, as shown in FIG. 12, the third injected fuel amount is gradually decreased in the low load region where the all-cylinder operation of the layer 3 is performed, and further decreased by one step at the load time point when shifting to cylinder deactivation. . As the load further decreases, the third injected fuel amount is increased.

  FIG. 13 shows the relationship between the total fuel amount and the load, which are the sum of the injected fuel amounts per cycle from the first to the third. As shown in FIG. 13, the total fuel amount per cycle of the first to third injected fuel amounts is increased at the time of load when shifting to cylinder deactivation. As a result, throttle loss in reduced-cylinder operation, that is, two-cylinder operation is suppressed, and stable SPCCI combustion can be realized while reducing the air-fuel ratio A / F even in a low load region that is disadvantageous for SPCCI combustion. it can.

  FIG. 14 shows an example of the relationship between the axial average effective pressure (equivalent to load) and the exhaust valve closing timing (EVC) in layer 2 and layer 3, and FIG. 15 shows the axial average effective pressure (equivalent to load) in layer 2 and layer 3. And an example of the relationship between intake valve opening timing (IVO). FIG. 16 shows an example of the relationship between the axial average effective pressure (corresponding to load) and the addition rate of external EGR in layer 2 and layer 3.

  As shown in FIG. 14, the closing timing after the top dead center in the layer 3 is about half as fast as the closing timing after the top dead center in the layer 2.

  On the other hand, as shown in FIG. 15, the intake valve opening timing is a constant value in both layer 2 and layer 3 including cylinder deactivation.

  Further, as shown in FIG. 16, the addition rate of the external EGR gradually decreases to the boundary region with the layer 3 in the layer 2. On the other hand, the layer 3 including cylinder deactivation is 0% in total, and therefore no external EGR is added.

(Control process for engine cylinder deactivation)
Next, the operation control of the engine 1 executed by the ECU 10 will be described with reference to the flowchart of FIG. First, in step S21 after the start, the ECU 10 shown in FIG. 4 reads the signals of the sensors SW1 to SW17, SW20 to SW24, SW31, SW51, SW101, SW102, and SW161. In the subsequent step S22, the ECU 10 determines the operating region of the engine 1. Here, whether or not the operation region is Layer 3 shown in FIG. 5, specifically, the low load region (1) -1 (SPCCIλ> 1 region), and if it is Layer 3, the temperature of the intake air And whether the temperature of the cooling water is equal to or higher than a predetermined value. If it is not Layer 3 and if it is Layer 3 and either one of the intake air temperature and the cooling water temperature is less than the predetermined value, the process returns to Step S21.

  If the operation region is layer 3 and both the intake air temperature and the cooling water temperature are equal to or higher than a predetermined value, the process proceeds to the next step S23.

  In step S23, it is determined whether or not it is a reduced cylinder operation (cylinder deactivation operation) region. As an example, whether or not it is a reduced-cylinder operation region is determined by whether or not the BMEP is 200 kPa or less. Specifically, for example, an accelerator opening sensor SW12 and / or a throttle opening sensor SW17 can be used.

  Here, if it is determined that the reduced-cylinder operation region, it proceeds to the next step S24. In step S24, as described above, the ECU 10 outputs a control signal to the swirl control valve 56 such that the opening degree is 0% to 15%. Thereby, the value of the swirl ratio becomes 4 or more and 6 or less.

  Next, a reduced cylinder operation is performed in step S26. That is, the ECU 10 outputs a control signal to the fuel supply system 61 so as to stop the supply of fuel to, for example, predetermined two cylinders out of four cylinders. Furthermore, for the other two cylinders that continue to operate, the ECU 10 performs a total of three injection timings shown in FIGS. 7 to 9 and the fuel amount at the injection timing corresponding to that shown in FIGS. Fuel is injected according to the above. Then, it returns to step 21.

(Other embodiments)
The technique disclosed here is not limited to being applied to the engine 1 having the above-described configuration. As the configuration of the engine 1, various configurations can be adopted.

  Further, the engine 1 may include a turbocharger instead of the mechanical supercharger 44.

1 Engine 10 ECU (control unit)
11 cylinders
17 Combustion chamber 171 Squish area 25 Spark plug (ignition part)
3 Piston 31 Cavity 312 Shallow bottom 401 Primary passage (first intake passage)
402 Secondary passage (second intake passage)
56 Swirl control valve (swirl generator)
6 Injector (fuel injection part)

Claims (6)

An engine having a plurality of cylinders, each having a combustion chamber;
An ignition part disposed in each combustion chamber;
A fuel injection section disposed facing each combustion chamber;
A controller that is connected to the ignition unit and the fuel injection unit and outputs a control signal to each of the ignition unit and the fuel injection unit;
When the engine is operated at a higher load than a predetermined load, the control unit supplies all of the cylinders with fuel to perform all-cylinder operation, while operating at the predetermined load or less. In order to perform the reduced-cylinder operation by supplying fuel to some of the plurality of cylinders,
In the reduced-cylinder operation, the control unit causes the fuel injection unit to inject fuel into a part of the plurality of cylinders, and controls the ignition unit to ignite the generated air-fuel mixture. Along with the air-fuel ratio larger than the theoretical air-fuel ratio, and the high compression ratio greater than a predetermined value,
The engine starts SI combustion due to flame propagation by ignition of the ignition unit, and then the unburned mixture undergoes CI combustion due to compression ignition ,
When the engine is operated below the predetermined load,
The control unit has a first combustion mode for performing SI combustion and CI combustion following the SI combustion at an air / fuel ratio equivalent to a value of the theoretical air / fuel ratio;
Switching between the second combustion mode for performing SI combustion and CI combustion following the SI combustion with an air / fuel ratio larger than the stoichiometric air / fuel ratio,
When the engine is operated in the first combustion mode, the compression ignition engine control device does not perform the reduced-cylinder operation . The control apparatus for a compression ignition engine according to claim 1,
Wherein, in the reduced-cylinder operation, the control device of a compression ignition type engine as the state in which also actuates the intake and exhaust valves of the remainder of the plurality of cylinders. The control apparatus for a compression ignition engine according to claim 1 or 2,
When the intake air temperature to the combustion chamber is lower than a predetermined value, the control unit is a control device for a compression ignition engine that performs all-cylinder operation instead of reduced-cylinder operation. In the control apparatus of the compression ignition type engine according to any one of claims 1 to 3,
When the coolant temperature of the engine is lower than a predetermined value, the control unit controls the compression ignition type engine to perform all-cylinder operation instead of reduced-cylinder operation. In the control device for the compression ignition engine according to any one of claims 1 to 4,
A swirl generator for generating a swirl flow in each of the combustion chambers;
In the reduced-cylinder operation, the control unit is a control device for a compression ignition engine that generates a swirl flow in the combustion chamber with respect to the swirl generating unit. The control apparatus for a compression ignition engine according to claim 5,
A control device for a compression ignition engine, wherein a value of a swirl ratio in the swirl flow is 4 or more.

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